Pump



June 12, 1956 c. o. WEISENBACH ET AL 2,749,844

PUMP

Filed Nov. 1 1951 4 Sheets-Sheet 1 I0 PILOTJ CO/VTEOL UPPLY INVENTOR. 3omens o. WE/JENBACH JAM/EL J FELZ J2. f%/MA/0 J C/GAL ATJ'UENE) June 12,1956 c, o. w s H ETAL 2,749,844

PUMP

4 Sheets-Sheet 2 Filed Nov. 1, 1951 z 3 Q g 3 m. Q NR W m 4v 5 a 0 Q Wm}47 4/ wk MWVWM I g Q fly Q m m Q \Q \ww NE, Q E c F h m wk Q r L W 3 mmw n3 g 3 %Q \A 1? s \N wk R w\\ 3 SQ ATTORNEY June 1956 c. o. WEISENBACHET AL 2,749,844

I PUMP Filed Nov. 1, 1951 4 Sheets-Sheet 5 T06 5057' M JEQT OFFSETQO/VTEOL PE FOECf 0 5% Q4]! PLATE A/vcus 414 INVENTOR. C/VAEMZS 0.wax/V540 ow/10,52 .2 F222 J6 freomw J C/G'AL ATTORNEY PUMP (Jharies O.Weisenbaeh, Samuel John Peiz, .lr., and Ferdimind I. Cigal, auth Feud,lnd., assignors to Bendix Aviation Corporation, South Bend, End, acorporation of Delaware Application November 1, 1951, Serial No. 254,262

2 Claims. (Cl. 103-162) This invention relates to pumps and moreparticularly to fuel pumps of the variable displacement, reciprocatingplunger, swash plate type for pressurizing fuel to prime movers, adaptedto respond automatically to a change in fuel flow and discharge pressurerequirements. Certain types of engines, notably turbojet engines foraircraft, require extremely high delivery pressures at the burnerdischarge nozzles to obtain good fuel atomization and efficientcombustion, and the nozzle pressure requirements may vary over a widerange between operation at sea level and at high altitudes. Thereciprocating plunger, swash plate type of pump is well adapted tosupply fuel at high pressures, but automatic regulation or variation ofpump delivery in relation to engine fuel requirements has heretoforeposed difiicult problems, solved only by means of relatively complexcontrol systems. it is therefore a primary object of this invention toprovide in a pump of the type specified a simple, compact andselfcontained regulator mechanism which acts to vary the angle of theswash plate as a function of pump discharge pressure.

A further object is to provide, in a pump of the type specified, simpleautomatic regulating means whereby great pump versatility is realized.

Another and more specific object is to provide, in a pump of the typespecified, automatic regulating means wherein the pump displacementvaries directly as a function of pump plunger inertia and inversely as afunction of the pump discharge pressure, and wherein a resilient meanstends to increase the pump displacement as an inverse function of pumpswash plate angle.

A further object of this invention is to provide a pump of the variabledisplacement type especially adaptable for use in the fuel supply systemof gas turbine engines for aircraft.

Other objects and advantages will be readily apparent from the followingdetailed description taken in connection with the accompanying drawingsin which:

Figure l is a diagrammatic view of a fuel system for an aircraft gasturbine engine;

Figure 2 is a right hand end view of one of the variable displacementpumps in Figure 1;

Figure 3 is a sectional view of one of the variable displacement pumpsshown in Figure 2 taken on line 3--3 of Figure 2 with a sectional viewof a throttling valve shown connected thereto;

Figure 4 is a left hand end view of the pump rotor and port insert,shown in section in Figure 3, with the pump plungers removed;

. Figure 5 is a right hand end view of the port insert, shown in sectionin Figure 3, with the ends of the rotor cylinders shown dotted and inthe same relation to the ports of said insert as is shown in Figures 3and 4-;

Figure 6 is a graph illustrating force versus pump cam plate anglecharacteristics for a particular control spring rate and at a constantR. P. M.;

Figure 7 is a graph illustrating the discharge pressure Patented June12, 1956 versus flow characteristics when a relatively small port insertrotation is used; and

Figure 8 is a graph illustrating the discharge pressure versus flowcharacteristics when a relatively large port insert rotation is used.

Referring to the drawings, and first to Figure 1, fuel from a fuelsupply tank (not shown) supplies the fuel system shown in Figure 1through conduit 10 to the burner nozzles (not shown) in the combustionchambers of a gas turbine engine for aircraft. The fuel supply is underautomatic regulation by means of a fuel control unit 12 which may be ofany type suitable for the function which it is designed to perform. Afuel control unit of the type illustrated in the copending applicationSerial No. 74,322, filed February 3, 1949, of Dray and Kuzmitz, nowPatent No. 2,628,472, may be used if the control amplifier unit, and itsassociated hydraulic circuitry, be deleted therefrom. The fuel ispressurized to the fuel control unit by means of variable displacementpumps, shown generally at 16 and 18, which receive fuel from inletconduits l0 and 19" and discharge fuel at high pressures throughdischarge conduits 2t? and 2% into common conduit 22. Interposed betweensaid pumps and fuel control unit 12 in conduit 22 is a throttling valveshown generally at 24 which maintains a constant head at all timesacross fuel control unit 12 and is sensitive to the control unit 12pressure drop between control inlet conduit 26 and metered fuel pressureconduits 28 and 30. A pressurizing valve and shut-off cock unit is showngenerally at 32 interposed between control unit l2 and the burnernozzles, not shown, in metered fuel conduit 28. The pressurizing valveproportions the fuel flow to the primary and secondary fuel manifolds,not shown, while the shut-off cock, when closed, enables the manifoldsto be drained to atmosphere While relieving the fuel pressure in conduit28 to the low pressure side of the fuel pumps through conduit 34. Theunit indicated at 36 is a barometric element designed to automaticallyvary the fuel feed as a function of changes in entering air density.This unit prevents the fuel-air ratio from becoming too rich as altitudeis gained thereby preventing excessively high burner temperatures andover-speeding of the engine.

Referring to Figures 2 and 3, the pump 16 comprises a rotor body 38arranged in a casing 40 and having a plurality of annularly disposedbores 42 which are angularly symmetrical with respect to the axis of therotor body, each of said bores containing a reciprocable hollowedplunger or piston, two of which are shown at 44 and 44'. The rotor body38 is journalled in bearings 46 and 48 along sleeve sections 50 and 52respectively of the rotor, said rotor being engine driven from drivingspline 54 which is connected to an internally splined drive 56 by shaft58. A nonrotatable assembly 59 consisting of a substantiallycylindrically-shaped control ring 60 and a swash or cam plate 61 iscentrally pivoted on a transverse axis thereof on trunnion pin 62. Arotatable auxiliary cam plate 64 is pivotal about a spring loaded thrustbushing 66, and has a series of holes, such as 68 and 68, therein whichare adapted to receive cup-shaped plunger or piston slippers 70. Theauxiliary cam plate 64 insures continuous sliding contact between camplate face 72 and piston slippers bearing surface 74 when the pump is inoperation. A universal jointure is formed between each piston slipper 70and the balled end 76 of each pump plunger. Springs 78 insure pumpplunger return after each pumping stroke and aid thrust bushing spring3%) in maintaining pressure contact between surfaces 82 and 34 of therotor and a port insert member 86 respectively. When the control ringand cam plate assembly 59 pivots on trunnion 62, as during a period ofchange in demand on the pump 16, the auxiliary cam plate 64 pivots aboutthrust bushing 66, said auxiliary cam plate being pivotally driven tofollow the movements of assembly 59 by the action of the pump plungers44 and 44' and the plunger slippers 70. A port insert 86 is held rigidlyagainst surface 88 of casing 40 by retainer- 90, bearing 48 and annularkey member 92. Arcuate-shaped inlet port 94 and outlet port 96 (Figures4 and in port insert member 86, offset with relation to the axis oftrunnion 62 as projected on port insert 86, and also offset in relationto the vertical axis of said member as shown in Figure 5, providecommunicating means between fuel inlet and outlet conduits 98 and 101),and rotor body cylinders 42. Filter means 102 is preferably inserted ininlet conduit 98, and a check valve 104 is included in discharge conduit100 to prevent reversal of flow downstream thereof.

A control compression spring 105 mounted between retainers 186 and 188in chamber 109, is opcra'ively con nected to the control ring and camplate assembly 59 by a control shaft 110 and a rod 112 pivoted at oneend to shaft 11(1 and at the other end to control ring 60. Spring 1135,which may be manually adjusted by a screw 114 in cover 116 and boss 118of the end plate 128 of the pump, tends to increase the angle of thecontrol ring and cam plate assembly 59.

An adjustment screw 126 limits the maximum pump flow by limiting themaximum angle to which the control ring and cam plate assembly 59 cango. A vapor vent 128 is connected to housing cavity 130 and a fuel vent132 connects cavity 130 to pump inlet pressure (connection not shown)thereby preventing fuel pressure buildup in cavity 130 due to leakagepast the pump plungers. A vent 133, connecting cavity 130 to chamber 109through control shaft 111 maintains pump inlet fuel pressure in saidchamber thereby insuring hydraulic balanie of said shaft. The driveshaft seals are indicated at 34.

Connected to pump discharge conduits and 22 is the automatic throttlingvalve shown diagrammatically at 24 in Figure 1 and in cross section inFigure 3. It functions to maintain a constant pressure drop across fuelcontrol unit 12, as hereinbefore mentioned, by area control of port 136.The valve itself, shown at 138, is shown as of cylindrical form mountedto slide in a sealed bushing 140 formed with a mounting flange 141,between which and a plate 142 is secured a diaphragm 144, the latterconstituting a movable wall between chambers 146 and 148. Chamber 146 isvented to chamber 148 by vent 150 and passage 152. A spring 154 has oneend abutting a fixed bushing 156 seated in a cap or cover 158, and theopposite end engaging a cup-shaped member 160 having a stem 162 slidablyprojecting through an opening 164 in the bushing 156, movement of themember 160 in a valve opening direction being limited by an adjustablenut 166. The valve 138 has connected thereto a stem 168 which projectsthrough the center of the diaphragm 144 and carries a cone-shapedabutment member 170 which projects into the cup 160. Metered fuelpressure in conduit 28 (Figure 1) is communicated to chamber 148 throughconduit 30, whereas throttling valve outlet pressure (fuel control inletpressure) is communicated to chamber 146 by means of passages 171 in thebody of valve 138, whereby the fuel pressure in conduit 26 will bemaintained at a substantially constant value over and above the fuelpressure in chamber 148 as determined by the effective force of spring154, or in other words, there will be a substantially constant pressuredrop across the fuel control unit 12 under all conditions of operation.

Referring now to Figures 4 and 5, rotor body bores 42 decrease in sizealong conical sections 172 (Figure 3), shown diagrammatically in Figure4, the plunger ends of the bores being indicated at 174 and the ports at176, said ports being connected to the cylinders by passages 177. Innerand outer circumferential boundaries of port insert 86 are indicated at178 and 179, respectively. To further clarify the relation of Figure 4to Figure 3, a cross section taken on line AA of Figure 4 would yieldthe view of the rotor body 38 and the port insert 86 seen in Figure 3.Fuel inlet port 94 of insert 86, which is curved to coincide with thepath of the revolving cylinder ports 176 during the intake stroke of theplungers, is spaced farther from the top dead center position(TDCposition of cylinder when plunger is at end of discharge stroke)than from the bottom dead center position (BDC-position of cylinder whenplunger is at end of intake stroke). The distance between the top deadcenter and port 94 in the direction of rotor rotation will normally begreater than the width of ports 176. In this construction, as thecylinders pass from top dead center to the point at which communicationis established between ports 94 and 176, the pressure in said cylindersremains at substantially fuel outlet pressure. The moment communicationbetween said ports is established the pressure in the respectivecylinders immediately becomes equal to the fuel inlet pressure inconduit 98. The high pressure in the cylinders passing from top deadcenter to the point of communication between ports 94 and 176 exerts aforce through the pump plungers urging cam plate 61 toward neutralposition. This force will increase and decrease as the pump outletpressure increases and decreases in response to variations in fueldemand. Fuel outlet port 96, which is curved to coincide with the pathof ports 176 during the discharge stroke, is spaced farther from bottomdead center than from top dead center. The distance between bottom deadcenter and port 96 in the direction of rotor rotation will normally begreater than the width of ports 176. Thus, as the cylinders pass frombottom dead center to the point at which communication is establishedbetween ports 96 and 176, the pressure in the cylinders remainssubstantially equal to fuel inlet pressure. Ports 94 and 96 arepreferably so disposed in relation to the bottom and top dead centerpositions, respectively, that they will be in communication with ports176 of the cylinders in bottom and top dead center positions. The spaceor portions of metal between top dead center and port 94 and betweenbottom dead center and port 96 may be varied from one pump to another tosatisfy the requirement of any particular installation or it may bevaried in any particular pump to meet special operating requirements byrotating insert 86.

Operation In the operation of the pump, drive shaft 58 rotates rotorbody 38 at some predetermined ratio of engine speed, whereby pumpplungers 44, 44 reciprocate in their respective cylinders as they arecaused to follow nonrotatable cam plate face 72 on the piston slippers70 by the combined action of springs 78, fuel inlet or discharge backpressure on the pump plungers, and auxiliary cam plate 64. As thecontrol ring and cam plate assembly 59 variably tilts, due to acombination of forces to be hereinafter described in detail, thedisplacement of the plungers 44, 44' varies, and therefore the fuel flowvaries, increasing as the cam plate angle increases relative to thevertical axis of trunnion pin 62, a maximum flow position as determinedby the setting of manual adjustment stop screw 126, and decreasing tozero as the vertical axis of the cam plate assembly becomesperpendicular to the axis of the rotor body.

In Figure 3 plunger 44, shown at bottom dead center, has just completedits fuel intake stroke having sucked fuel from conduits 10' and 98through inlet port 94 and is ready to begin discharging fuel at highpressure through discharge port 96. Plunger 44, shown just prior toreaching top dead center, is shown completing its discharge stroke andwill remain substantially at pump discharge pressure until it reachesthe approximate position past top dead center presently occupied by theplunger immediately adjacent to it in the direction of rotation (Figures4 and 5).

Three distinct primary forces are present which tend to increase theangle of the control ring and cam plate assembly 59 when the pump ispumping fluid. Firstly, the

control spring 105 exerts a force in a cam angle increasing direction asdo the springs 78 of the pump plungers which are moving across the lowerhalf of the cam plate. Secondly, a relatively large motional force ofinertia is acting on the cam plate to increase its angle as a result ofthe change in direction of movement of the pump plungers as they passthrough the plane of bottom dead center just after termination of thesuction stroke. This inertial force varies in a direct linear relationto the cam plate angle and directly as the square of the pump R. P. M.In addition to these primary forces the relatively small force of inletfuel pressure, acting on the plungers travers ing the lower half of thecam plate, acts in a cam plate angle increasing direction, the fuel atinlet pressure contained within casing 40, due to the venting of thecasing cavity through port 132 to the inlet side of the pump, having azero resultant force about trunnion 62.

Two primary forces are present which tend to return the cam plateassembly to a no-fiow or neutral position. Firstly, the plunger springs78, which are in a condition of maximum compression at top dead center,tend to return the cam plate to neutral while the plungers aretraversing the top half of the cam plate. Secondly, the carry-overeffect of pump discharge pressure, due to the offset portingarrangement, results in a large force which tends to return the camplate to a neutral position. For example, as best illustrated in Figures4 and 5, a plunger passing top dead center at pump discharge pressureremains substantially at said pressure until vented to inlet port 94thereby exerting a neutralizing force on the cam plate 61 which is equalto discharge pressure multiplied by the effective area of the plungertimes the moment arm of the cam plate about trunnion 62. This carry-overeffect may be varied as required by varying the port insert offsetangle. On the other hand, a plunger passing bottom dead center remainssubstantially at pump inlet pressure until such time as it breaks intothe discharge port 96.

Referring now to Figure 6, the characteristic curves of those forcestending to increase the cam plate angle are shown. it is seen that atany given R. P. M. the force of inertia increases linearly withincreasing cam plate angle and that the control spring force decreaseslinearly with increasing cam plate angle at a rate greater than theincrease in inertia force so that a resultant negatively-sloped curve isinsured. It has been found that, for best pump stability, a resultantcam plate angle increasing force should be used which decreasesappreciably with increase in the angle of the cam plate. The controlspring force and the pistons spring forces are, of course, substantiallyconstant for all pump R. P. M.s while, as previously mentioned, theforce of inertia varies as the square of the R. P. M. Therefore, theresultant force curve shown in Figure 6 will, as the R. P. M. increasesand decreases, decrease and increase its negative slope respectively, sothat a fan of resultant force curves all emanating from the same pointon the coordinate of force results. At each R. P. M. and cam plate anglethe corresponding resultant force tending to increase the cam plateangle is balanced by the forces tending to return the cam plate to aneutral position. This balance of cam plate angle increasing forcesautomatically results from the action of the throttling valve 24-(Figure 2) and the effect of the pump discharge pressure on cam plateposition as determined by the port insert offset angle. As, for example,the pilot moves lever 14 (Figure 1) to an increased power setting, thefuel flow to the engine increases due to the action of fuel control unit12, resulting in an immediate decrease in fuel pressure in chamber 146of throttling valve 24 and an immediate increase in fuel pressure inchamber 14% of said valve. As a result, valve 138 opens resulting in amomentary decrease of pump discharge pressure whereby the forces tendingto increase the angle of cam plate 61 overcome the forces acting todecrease said angle and the cam plate moves to a new position resultingin the desired increase in fuel flow to the fuel control unit 12 as thenew discharge pressure acts, due to the porting arrangement, to balancethe decreased control spring force and the increased force of inertia.

Referring now to the characteristic pump curves shown in Figures 7 and8, it is to be noted that the same control spring rate is used withreference to each set of curves, the differences in curve contour at anygiven R. P. M. being solely due to variation in the angle of port insertoffset. The characteristic curves at the indicated R. P. M.s in Figure 7result from the use of a small port insert offset whereas those shown inFigure 8 result from the use of relatively large port insert offset. Tofurther clarify the differences in pump operation as illustrated bythese curves a comparison of the 2500 R. P. M. curves in each case willbe made. With no restriction in the discharge con duits 20 and 22 thepump discharges fuel at a rate indicated by points a and a in Figures 7and 8 respectively, at which points the cam plate angle is maximum. Asthe discharge pressure is increased from atmospheric, as by means of theaction of a valve such as that shown at 24, it follows curves a'b'c' andabc, in each case. From a to b and from a to b the pump control is in aposition of maximum displacement whereas from b to c and from b to 0,pump displacement is steadily decreasing due to the action of dischargepressure carry-over effect on the port insert which returns the camplate to neutral at points c and 0 respectively, at which points thedischarge line restriction is fully closed and no fuel is flowing. Thebreaking of the curves at points I) and b is a result of the dischargepressure action overcoming the control spring force and the force ofinertia at maximum cam plate angle, and the lower pressure at whichcurve abc breaks is due to the greater port insert offset angle. Curveportions ab and a'b would be vertical but for fuel leakage past the pumpplungers as discharge pressure increases. The upward negative slopes ofcurve portions bc and [1c are due to the control spring rate, which maybe varied as desired. Point d of Figure 8 represents a theoretical pointof intersection of the famiiy of curves of Figure 8 if there were nopiston leakage to the inner cavity of pump casing 40. The decreasingnegative slope with increasing R. P. M. of the curves of Figures 7 and 8respectively, after the respective points of breaking, is caused by theincreasing inertial force with increasing R. P. M., so that a curvedrawn through the respective break points of the curves of Figure 7, forexample, would be, theoretically, a s uare curve. It is seen that theportion of the 3500 R. P. M. curve of Figure 7 after the break pointassumes a positive slope due to a greater decrease in inertial forcesthat increase in control spring force as pump cam plate angle decreases(Figure 6), which is undesirable, and that therefore a greater portinsert offset, such as used with Figure 8, will result in better pumpstability at that speed. Desired variations in the characteristic curvesof the pump can be easily obtained with this invention by simplychanging the control spring rate used and/or the port insert offset. Ithas been found that, with proper selection of a control spring andoptimum port insert offset, the stability of this type of pump isgreatly improved as is the response time to changes in engine operatingconditions. It has also been found that the subject pump control systemis ideal for pumps operating in tandem (Figure 1), particularly whereequal distribution of pump load is desirable under all conditions ofoperation.

While we have herein shown and described a preferred embodiment of ourinvention, it will be evident that the invention may be embodied inother forms. For example, the rotor body, as described in the preferredembodiment, may be made the stator with the engine drive connected tothe swash plate and port insert, in which case a non rigid connectionbetween the control spring shaft and the swash plate would be used.Likewise, rotor body cylinders may be parallel to the axis of the pump.Also, the swash plate, as shown in the preferred embodiment, may beaxially offset with respect to the trunnion, in either direction,thereby providing a larger or smaller moment arm on which the reactiveforces of pump discharge pressure may act and in connection with which asmaller or larger port insert offset angle would be used respectively.Other modifications will likewise be apparent to those skilled in theart from the foregoing description taken in connection with theaccompaying drawings.

We claim:

1. In a pump, an inlet and an outlet port, a plurality of cylindersadapted to register with said ports, a piston in each of said cylinders,a thrust member formed to engage said pistons, said member beingcentrally pivoted on a transverse axis thereof, resilient meansconnected to said member on one side of a plane through said axis forurging said member in one direction about said axis, said outlet porthaving one part disposed on said one side of said plane and having asecond larger part disposed on the other side of said plane whereby ahydraulic reaction force is directed against a portion of said pistonsto urge said member in a direction opposite to said one direction.

2. In a pump, an inlet and an outlet port, a plurality of cylindersadapted to register with said ports, a piston in each of said cylinders,a thrust member adjacent said cylinders to control the movement of saidpistons, a pivot axis for said member passing substantially through thecenter thereof, resilient means connected to said member on one side ofa plane through said axis for urging said member in a direction toincrease piston stroke, at least amajor portion of said outlet portbeing located on the side of said plane opposite to said one sidewhereby the reaction force against said pistons urges said member inadirection to decrease piston stroke.

References Cited in the file of this patent UNITED STATES PATENTS1,466,092 Egersdorfer Aug. 28, 1923 1,785,356 Lawser Dec. 16, 19302,288,768 Zimmermann July 17, 1942 2,299,233 Hoffer Oct. 20, 19422,299,234 Snader et a1 Oct. 20, 1942 2,517,313 Hooker et a1. Aug. 1,1950 2,546,583 ,Born Mar. 27, 1951 2,628,472 Dray et a1 Feb. 17, 1953FOREIGN PATENTS 571,622 France Mar. 27, 1951

